This invention comprises an improvement on the bearings systems of U.S. Pat. No. 6,877,901 B2, dated Apr. 12, 2005, and U.S. Pat. No. 7,025,579 B2, dated Apr. 11, 2006.
Turbochargers have been used on both diesel and gasoline engines to increase power output, reduce fuel consumption, and compensate for high altitude power loss for many years. The success of the small turbocharger to a large part can be attributed to the development of satisfactory bearing systems that allow the machines to operate at very high speed and attain sufficient durability to make them commercially viable products.
Early small turbochargers employed stationary sleeve bearings and rotating assemblies with large diameter stiff shafts where the first critical speed was higher than the maximum operating speed of the machine. Since the maximum pressure output of the early turbocharger compressors was of the order of 1.6 times atmospheric pressure, the bearing friction losses of the large diameter bearings were acceptable.
When the potential of turbocharging high-speed diesel and gasoline engines became obvious, the pressure output of the turbocharger compressors needed to be increased substantially and the friction losses of the large diameter shafts became unacceptable.
When the shaft sizes of the small turbochargers were reduced to lower the friction losses, the critical speed of the rotating assembly fell within the operating speed range of the turbocharger and bearings systems had to be designed and developed to permit the rotating assemblies to pass through their first critical speed without failure. This pioneering work, performed primarily in the 1960's, resulted in the development of small diameter floating sleeve bearing systems where the sleeve bearing was allowed to rotate at a fraction of the speed of the rotor assembly, thereby reducing the friction losses to acceptable values.
The floating sleeve bearings still in use currently in commercial turbochargers incorporate an inner and outer oil film where the outer oil film provides a damping cushion that permits the rotor to pass through its critical speed without reaching a vibration amplitude that would cause failure of the bearings. The floating sleeve bearings also permit minor radial movement of the rotor allowing it to find and rotate about its center of mass, thereby eliminating radial forces that would be generated if the rotor were constrained to rotate about its geometric center.
The inner and outer oil films provide the necessary lubrication to prevent wear and also provide a cushion against shock and vibration. Examples of these successful bearing systems are illustrated in U.S. Pat. Nos. 3,056,634, 3,096,126, 3,390,926, and 3,993,370.
In the floating sleeve bearing systems described above, it was necessary to provide a thrust bearing to carry the axial loads imposed by the actions of the turbine and compressor wheels used in turbochargers, and a collar was provided on the rotating shaft to bear against stationary thrust members. Since the friction loss in radial thrust bearings is a function of the fourth power of the radius, the collar attached to the shaft causes a relatively high thrust bearing friction loss which, when added to the friction loss of the sleeve bearings, results in a substantial total frictional loss for the complete bearing systems.
It is advantageous to have a bearing system in small turbochargers that has a high mechanical efficiency due to the very high speed of rotation of the rotor assemblies. Since sleeve bearing losses are proportional to the square of the rotational speed of the shaft, there have been numerous attempts in the past to develop systems that use ball bearings in small turbochargers. One such system is described in U.S. Pat. No. 4,370,106 that discloses the use of a non-rotating, elongated cylinder with a ball bearing mounted in one end and a sleeve bearing at the opposite end. The elongated cylinder is prevented from rotating by a square flange on the end carrying the ball bearing; the square flange engaging stops in a stationary housing member. Lubricating oil is provided around the outside diameter of the elongated cylinder to provide a damping cushion for orbital motion of the rotating assembly caused by residual unbalance. This residual unbalance forces the non-rotating cylinder to orbit radially at a very high frequency and causes the mating surfaces of the square retaining flange to be subject to fretting. Thus, a solid lubricant pad was used between the mating surfaces of the square flange and corresponding stationary housing surface to mitigate the fretting problem, however, this problem still contributed to a limited service life in turbochargers using this bearing system.
The pursuit of satisfactory ball bearing systems for small turbochargers continued with the emergence of the system shown in U.S. Pat. No. 4,641,977. In this system, a ball bearing is mounted in one end of an elongated cylinder and includes a separate floating sleeve bearing near the opposite end. The elongated cylinder incorporates a round, radially extending flange on one end and the cylinder is free to rotate at a fraction of the speed of the shaft. The radially extending flange on the end of the rotatable cylinder engages corresponding surfaces of the stationary housing and carries thrust loads of the rotating assembly in both directions. Friction losses with this system are reduced due to the ball bearing and floating sleeve bearing resulting in a higher mechanical efficiency than previous systems; however, since the system contains one sleeve bearing, there remained room for improvement in mechanical efficiency if a double ball bearing system could be developed.
This invention provides improvements over the bearing systems disclosed in U.S. Pat. Nos. 6,877,901 B2, and 7,025,579 B2. Each of these prior bearing systems employ two ball bearings mounted in an elongated cylinder that incorporates a radially extended annular flange on one end to carry thrust loads. The inner races of the ball bearings are separated by an elongated spacer, and the inner races and spacer are clamped together to rotate as an integral part of the rotor assembly.
The bearing system disclosed in U.S. Pat. No. 6,877,901 B2 shows an angular contact ball bearing mounted in both ends of a rotatable elongated cylinder. The cylinder is supplied with lube oil over its outer diameter that provides a cushion against shock and vibration and lube oil is admitted to the inner diameter to provide lubrication and cooling for the angular contact ball bearings. Under normal operating conditions, the system functions satisfactorily, however, under extreme conditions where engine exhaust temperatures reach very high levels, the turbine end of the shaft experiences axial expansion that can cause the inner race of the turbine end bearing to move axially away from the outer race that is pressed into the elongated cylinder and allows the balls in the angular contact ball bearing to skid in the races.
The bearing system disclosed in U.S. Pat. No. 7,025,579 B2 shows angular contact ball bearings mounted in both ends of an elongated cylinder that is constrained against rotation by two elastomeric members placed between the elongated cylinder and the corresponding mounting bore in the bearing housing. The elastomeric members allow small radial orbital movement of the elongated cylinder caused by rotor residual unbalance to take place, and also provide a cushion against shock and vibration. The ball bearings in this system are grease-lubricated and do not require a supply of pressurized lubricating oil from the engine.
The outer diameter of the non-rotatable elongated cylinder is exposed to a supply of cooling fluid from the bearing housing fluid cavity, which serves to carry away heat generated in the bearings when operated at high speed. The inner races of the ball bearings are separated by an elongated spacer, all of which are clamped on the shaft and rotate with it. High exhaust temperature during extreme conditions of operation causes the shaft to expand axially, and the inner race of the turbine end bearing moves axially in accordance with this expansion. This movement of the inner race of the turbine end bearing can cause the balls in the bearing to skid in the races.